Rotary engine with non-uniform piston speed.

Field of Invention

This invention concerns rotary internal combustion engines of the "scissor action" type which are based on a circular cross section (annular section) cylinder which houses one pair of twin (or more) vane-bearing cylindrical portions concentrically mounted which rotate independently of each other at differential (out of phase) speeds in order to define between them through each full rotation increasing and decreasing volumes necessary for the intake, compression, expansion and exhaust cycles of the engine.

Background to the Invention

It has been proposed inter alia in UK Patent Specification No. 2007771A to generate rotational energy by means of two independently rotatable rotary piston pairs 10', 10" and 12', 12" which are mounted coaxially but can rotate independently of each other. The two coaxial rotary piston pairs make complete rotations whilst alternately accelerating and decelerating, the net force being obtained by creating a resistance to rotation of one rotary piston during the compression and firing cycle associated with the other and then reversing the roles of the two pistons. In this way each piston accelerates and t decelerates continuously as it rotates out of phase with the other piston.

In the engine described in UK 2007771A the two rotary piston pairs are mounted for separate coaxial rotation, •

with ^' each piston pair forming a two bladed propeller when viewed end on, the two sets of pistons defining four variable chambers therebetween which will allow four power strokes per 360° revolution, the engine working on a four stroke principle (i.e., intake, compression, expansion, exhaust) .

The main characteristic of such an engine i-s that the motion of the two output shafts is not constant speed rotation but instead is made up of two accelerations separated by two decelerations during each 360" degrees of rotation with the acceleration and deceleration pattern of the one shaft out of phase with that of the other. Such a motion results from the so-called hesitating progression of the piston pairs around the cylinder.

In UK 200777'lA the -pulsating .cyclical motion of the two output shafts produced by this hesitating progression of the pistons, is converted into constant speed rotation using two pairs of continuously meshing elliptical gears • 36, 40 and 34, 38 , the one pair being driven by one output shaft 20 and the other by the other output shaft ^' 15. The second of the two elliptical gears of each pair ie 38 and 40 is keyed to. a common output shaft 30 and mounted thereon 180° out of phase. This permanent phase differential between the two driven elliptical gears 38 and 40 maintains a corresponding phase reversal between the earlier gears in each gear train driven by the two output shafts 18 and 20 and thereby causes the relatively rotatable rotary piston pairs to remain 180° out of phase in terms of acceleration and deceleration (so that whilst one is speeding up the other will be slowing down and vice versa) .

The construction of such a rotary engine is in practice rendered very difficult and expensive due to the requirement for constantly meshing elliptical gears (36, 40 and 34, 38) capable of transmitting power. The tooth angle of such gears varies continually around the gear wheel and in general, it has proved impossible to construct such gear wheels which do not suffer very rapid wear during use and the development of such a drive unit has therefore been arrested.

In this design each piston assembly is interlocked via the meshing toothed wheels to the other piston assembly. Although the prior published specification attempts to imply that any form of gears can be used, provided they are mounted eccentrically, it is in practice impossible to achieve the arrangement shown with any gears' other than elliptical gears.

It is an object of the present invention to provide an improved rotary engine.

Summary of the Invention

According to the present invention, there is provided a rotary internal combustion engine in which two vane- bearing members rotate in a common cylinder with varying relative velocities so that the chambers defined between any two vanes vary in volume, the rotation of the two vane-bearing members driving a common rotary output shaft through a drive mechanism such that both vane-bearing members alternately accelerate and decelerate with respect to the common rotary output shaft but the said accelerations and decelerations of one vane-bearing member are out of phase with those of the other by a

predetermined phase angle, characterised in that the drive mechanism accommodates the said alternate rotational acceleration and deceleration of the vane- bearing members with respect to the rotary output shaft at least in part by the action of one or more members of the drive mechanism having a round cross-section and rotating in continued revolutions about centres other than their own geometric centres and which further interact with other rotary members of the drive mechanism.

In a typical engine there will be only two vane-bearing members, and the phase angle will be 180°. However, the present invention encompasses arrangements in which there are more than two vane-bearing members and the phase angle is other than 180°. For example, there could be three vane-bearing members with mutual phase angles of 120".

In one arrangement embodying the present invention, the drive mechanism comprises, in respect of a first -vane- bearing member, a first rotary unit mounted on a first shaft to rotate with the first vane-bearing member and a second rotary unit mounted on a second shaft to rotate with the common rotary output shaft, the axes of the first and second shafts being parallel but laterally displaced with respect to one another and the first and second rotary units rotating face-to-face in spaced-apart parallel planes, the first and second rotary units being joined by a transmission member (pin) extending between them and being mounted slidably in a radial slot in one at least of the first or second rotary units and being mounted in a fixed position on the other rotary unit so that relative rotational movement of the opposed faces of the first and

second rotary units induced due to the lateral displacement between their respective axes of rotation is accommodated by a combined rotational and sliding movement of the transmission member (pin) of one rotary unit in the radial slot of the other rotary unit,

and in respect of a second vane-bearing member, a first rotary unit mounted on a first shaft to rotate with the second vane-bearing member and a second rotary unit mounted on a second shaft to rotate with the common rotary output shaft, the axes of the first and second shafts being parallel but laterally displaced with respect to one another and the first and second rotary units rotating face-to-face in spaced-apart parallel planes, the first and second rotary units being joined by a transmission member (pin) extending beween- them and being- mounted slidably in a radial slot in one ^' at least of the first or second rotary units and being mounted in a fixed position on the other rotary unit so that relative rotational movement of the opposed faces of the first and second rotary units induced due to the lateral displacement between their respective axes of rotation is accommodated by a combined rotational and sliding movement of the transmission member (pin) of one rotary unit in the radial slot of the other rotary unit.

The second rotary units of each vane-bearing member could have a common second shaft geared to rotate with the common rotary output shaft and the pin or radial slot of the second rotary unit of the first vane-bearing member is 180" out of phase with the pin or radial slot of the second rotary unit of the second vane-bearing member.

In an alternative arrangement embodying the present

invention, the drive mechanism comprises, in respect of each vane-bearing member, a first circular gear mounted eccentrically on a first shaft, a second circular gear meshing with the first and mounted eccentrically on a second shaft, a third circular gear mounted concentrically on the second shaft and a fourth circular gear meshing with the third and mounted concent ically on a third shaft, the second shaft being mounted so that its axis can be laterally displaced over an arc centred on the axis of the third shaft.

Preferably, each respective first shaft is constrained to rotate with the respective vane-bearing member and each respective* third shaft is or is constrained to rotate with the rotary output shaft.

*

In one embodiment of the present invention, a rotary engine of the type described in which two pairs of eccent ically mounted gears are maintained in constant meshing engagement, and in which the driven gears in each of the said pairs are keyed separately to a pair of stub shafts, has the features: (1) the meshing gears are circular; (2) the stub shafts are mounted for rotation in bearings which allow relative lateral movement of the stub shafts to accommodate the eccentric movement of the driven gear wheels; (3) said stub shafts having a conventional circular gear wheel, concentrically keyed on each stub shaft, for meshing with one of a pair of output gears (which are similar conventional circular gears keyed to a common output shaft and mesh continuously with the circular gears mounted concentrically on the two stub shafts), and. (4) means constraining the lateral oscillatory movement of the stub shafts to follow a segmental arc of a curve having a centre of curvature

coincident with the axis of rotation of the said common output shaft.

In another embodiment of the present invention, in a rotary internal combustion engine there is provided a circular cross section, or annular, or other cross section cylinder in which rotate two independently rotatable coaxially mounted members each having at least two piston vanes, one of such two members being coupled to the first of a pair of two continuously meshing conventional circular gears, the two gears of each such pair being eccentrically mounted respectively on a pair of stub shafts, one of each such pair of stub shafts being mounted for rotation in fixed bearings while the other is mounted for rotation in bearings which allow lateral oscillatory displacement thereof in order to accommodate shaft displacement induced by the eccentric rotation, each such laterally displaceable stub shaft having a conventional circular gear wheel concentrically keyed thereon which in turn continuously meshes with one of a pair of output gears which are conventional circular gears coaxially keyed to a common output shaft, the lateral oscillatory movement of each oscillating stub shaft being constrained to follow a segmental arc of a curve having a centre of curvature coincident with the axis of rotation of the said common output shaft, and with the second of the output gears of the common output shaft continuously meshing with the second pair of the two eccentrically rotating continuously meshing circular gears which is coupled to the other rotatable member so that an even rate of rotation of the common output shaft will cause a hesitating progression of the rotatable members and their respective piston vanes about their axis of rotation.

Accommodation of the lateral displacements due to the eccentrically mounted gears may be achieved by permitting both stub shafts to oscillate laterally in a controlled manner to maintain the continuous meshing.

Meshing of the eccentric circular gear pairs may be perfectly aligned by means of connecting members which maintain a constant distance between the geometric centres of the meshing eccentric circular gear pairs whilst allowing their centres of rotation to be displaced laterally.

The connecting members are carried by rotatable rings, themselves freely rotatable by means of bearing assemblies about the geometric centres of the eccentric meshing gears .

Likewise the concentrically keyed gear wheel on each stub shaft and its associated output gear wheel are preferably kept in meshing alignment by means of similar connecting members (or keepers) which maintain a constant distance between the geometric centres of the meshing gears, whilst allowing the stub shaft to move laterally about the said segmental arc of curvature centered on the output shaft axis .

The cylinder for the rotary piston vanes may be formed at one end of a housing and in the drive mechanism located at the other end thereof.

Alternatively the part of the drive mechanism associated with one of the piston vane pairs is located on one side of the cylinder containing the rotary pistons whilst the part of the drive mechanism associated with the other of

the piston vane pairs is located on the other side of the said cylinder.

In a development of the invention, the common output shaft may serve as the common output shaft of further rotary piston power units. Typically there are two, four or eight power units in all, although eight is not a maximum number by any means.

Balancing masses are conveniently located on the eccentric gears and on the shafts, particularly the stub shafts, to balance the transmission unit and smooth out unwanted out of balance forces brought about by the eccentrically rotating (circular) gears.

Where the piston vanes are arranged in pairs, a gearing ratio of 2:1- is required between the piston vanes and the eccentric circular gear shafts to provide for a four stroke cycle. Other gearing ratios can be used to provide other cycles, such as two stroke. If more than two vane- bearing members are used, or each vane-bearing member bears other than two vanes, different gearing ratios may be required.

It is to be understood that although the piston vanes have been described as being arranged in pairs, (and in consequence.180 ° apart) more than two piston vanes can be affixed to each of the independently rotatable parts of the engine. Thus for example three (or more) vanes may be arranged at 120° (or less) apart respectively and interposed with a corresponding number of piston vanes on the other rotatable part of the engine. Where more than two such piston vanes are ^' so located, it is of course necessary to alter the ratio of the gearing between the

vanes and the ^' eccentric circular gearshafts so as to still provide for a true four stroke cycle as far as each "piston" is concerned.

In order to reduce -unwanted out of balance forces, the concentric toothed gears mounted on the stub shafts and output shafts may be duplicated, with one pair of concentric meshing gears on one side of an assembly of eccentric gears and the duplicate pair on the other side the said assembly of eccentric gears.

The full advantage of this configuration is realised if the two meshing concentric toothed gears of each said pair are kept at a constant. spacing between each other by means of a keeper of the type previously mentioned whilst still being able to oscillate about the axis of the output shaft to accommodate lateral movement of the stub shaft.

Brief Description of the Drawings

The invention will be now be described by way of example with reference to the accompanying drawings in which:

Figure 1 is a part cutaway exploded perspective view of a rotary piston engine power unit of the type described herein;

Figures 2(a) to 2(h) show the different parts of the cycle as the rotary pistons proceed to describe the complete 360" revolution;

Figure 3 is a diagramatic view from one side of a transmission drive unit for use with the rotary piston power unit such as shown in Figure 1, to form a complete

rotary engine embodying the present invention;

Figure 4 is a similar side view to that of Figure 3 but with the eccentric gears at a different angular position of rotation;

Figure 5 is a plan view of a gear train assembly of which Figures 3 and 4 are diagrammatic views;

Figure 6 illustrates an alternative arrangement for securing eccentrically mounted gears in continuous meshing engagement;

Figure 7 is a plan view similar to Figure 5 of an alternative rotary piston power unit in which the "eccentric" mes.hing gears are located on oppsite sides, of a -central piston containing housing;-

Figure 8 is a plan view analogous to the views of Figure 5 and Figure 7 , of a further transmission drive unit for use with the rotary piston power unit such as shown in Figure 1, to form a complete rotary engine embodying the present inven ion;

Figure 9 is a side view of one pair of rotary units, viewed in the direction of the arrow IX in Figure 8;

Figure 10 is a side view of the other pair of rotary units, viewed in the direction of the arrow X in Figure 8; and

Figure 11 is a view from above of the rotary unit 165 in section along line Xl-Xl in Figure 10.

Description of preferred embodiments

In Figure 1, which is an exploded perspective view of the component parts of a rotary piston power unit, a first pair of radially opposed pistons 10 and 12 are shown carried by a cylindrical member 14 having a bore 16, through which a shaft 18 can extend. The latter is an axial extension of a cylindrical member 20, similar to the member 14, and carries a further two radially opposed pistons 22 and 24, corresponding to 10 and 12.

With the shaft 18 in the bore 16, the two assemblies can be moved together, so that the end 26 of the cylindrical member 14 abuts the end 28 of the cylindrical member 20 and the radial pistons 10, 12, 22 and 24 are circularly spaced around the axis of rotation of the assembly.

As will be seen from Figure 2 which illustrates diagrammatically a known engine cycle, the pistons are located within a cylindrical housing 23 having in it an ignition point 25 with exhaust ports 27 and inlet ports 29 associated therewith.

Net drive is achieved by resisting the rotary motion of shaft 18 whilst permitting unit 14 to rotate and causing an explosion to occur between radial piston 24 and radial piston 10. The expansion resulting from the explosion causes the unit 14 and radial pistons 10 and 12 to rotate relative to unit 20 and pistons 22 and 24.

A second firing stroke is achieved by then introducing a resitance to the rotation of unit 14 whilst allowing unit 20 to accelerate and causing a further explosion this time between, for example, 12 and 24, so that 24 is

accelerated in the same sense as 10 was, this time relative to 12.

Although not shown, sealing is provided to ensure that exploding gases are constrained between the respective piston faces.

It will be seen that the resulting drive appears on the axial continuation of the cylindrical member 14 and the shaft 18. Both rotate in the same direction but cyclically accelerate and decelerate during each revolution. Whilst the one is accelerating the other is decelerating and vice' versa.

It is not possible to combine the two outputs and drive a common shaft without a complex transmission unit since the latter is required to introduce the cyclic resistance to rotation and alternate decelerations of the rotating units.

The invention is concerned with the provision of an improved transmission unit of this type for use in such an engine, and details of an embodiment of this new unit are found in Figures 3, 4 and 5.

In Figure 3, the cylindrical member 14 is shown keyed to a large diameter toothed wheel 30, which meshes with and drives a smaller diameter pinion 32. The pinion 32 has half the number of teeth as does 30. 32 is likewise keyed to a shaft 34, which itself is keyed to a first eccentric toothed wheel 36, mounted for rotation about an eccentric axis corresponding to shaft 34. The toothed wheel 36 meshes with an identical -wheel 38 also mounted for eccentric rotation and keyed to a shaft 40.

Although not shown in detail in Figures 3 and 4, means is provided for maintaining the distance between the two actual centres of the toothed wheels 36 and 38 at a constant distance apart. One form of such device is described in more detail with reference to item 70 in Figure 5. The actual centre of toothed wheel 36 is shown at 42 and the actual centre of wheel 38 at 44. The mechanism for maintaining these two toothed wheels at the constant distance apart which is described with reference to Figure 5 essentially comprises a thin housing 70 containing bearings permitting free rotation of the two wheels 36 and 38 about their actual centres relative to the housing. The shaft 40 is carried in a sliding bearing made up to two sliding parts 46 and 48, which move in a slideway 50, which is a segmental arc ^' having a mean radius of curvature centred on an axis 52. The latter is coincident with the axis of rotation of an output shaft 54 and drive is transmitted to the latter via a pair of identical meshing gears 56 and 58, the former keyed concentrically to the shaft 40 and the latter to the shaft 54.

It will be seen that as the pair of eccentrically mounted meshing gears 36 and 38 rotate, the position of the shaft 40 will move in space. The path of its movement will follow the arc of the slideway 50. Therefore, whatever position is adopted by the shaft 40, the gear 56 will always remain in constant mesh with the gear 58 enabling rotational drive to be transmitted thereto.

The slidway 50 can be considered to be constraining the path of movement of the shaft 40. However, as shown in Figures 3, 4 and 5, a further thin ^" housing 61 can be

provided.to maintain the centres of concentrically mounted gears 56, 58 at a constant distance, and this will also constrain the movement of the shaft 40 to an arc centred on the axis 52. Therefore it is not strictly necessary to provide the slideway 50 when the housing 61 is present. A person skilled in the art will appreciate that there are various possible constructions which will constrain the shaft 40 in the necessary manner.

Figure 4 is a similar view of the transmission unit after toothed wheel 30 has rotated in an anticlockwise manner (about 25°) from the position shown in Figure 3, thereby rotating pinion 32 and toothed wheel 36 in a clockwise manner, causing the actual centre 42 of the toothed wheel 36 to precess around the eccentric axis of notation (the shaft 34) of the toothed wheel 36.

By virtue of the meshing engagement of 36 and 38, the latter will rotate in an anticlockwise sense about its eccentric axis of rotation (shaft 40) and the eccentric throw of the meshing gears is accompanied by the sliding of the bearing 46 and 48 in the slideway 50.

The rotation of shaft 40 (due to the drive transmitted through the meshing eccentrically mounted gears 36 and 38) causes toothed gears 56, 57 and 56" , 57' (see Figure 5) also to rotate in an anticlockwise sense, thereby causing final drive gears 58, 59 and 58', 59' (see Figure 5) to rotate in a clockwise sense. The displacement of the shaft 40 moves the centre of rotation of the driving wheel 56, 57 and 56', 57' (see Figure 5) but i-n view of the arcuate shape of the slideway 50 the pairs of wheels 56, 58 (and 56*, 58' ; 57, 59 and 57' , 59' see Figure 5) remain in constant meshing engagement. Drive is therefore

continuously transmitted to the output shaft.

The arrangement diagrammatically shown in Figures 3 and 4 is one half of the overall arrangement which can be seen in Figure 5. Here the cylindrical member 26 and ^' inner shaft 18 are shown communicating with toothed wheels 30 and 30'. The same reference numerals are used to denote similar components in the two transmission trains and the same reference numerals are used in Figure 5 as have been used in Figures 3 and 4 for identifying components.

The large diameter toothed wheel 30 is matched by the wheel 30' attached to the shaft 18 and both drive smaller diameter toothed pinions 32 and 32' having half the number of teeth of wheels 30 and 30' as shown in Figure 5. The latter are attached to stub axles 34 and 34' respectively.

The arrangement shown in Figure 5 uses a pair of toothed eccentric wheels in place of a single eccentric wheel 36 and these are shown at 36 and 37. Likewise these mesh with similar eccentric wheels 38 and 39.

In the case of the transmission drive from the shaft 18, eccentric wheels 36' and 37' mesh with eccentric wheels 38' and 39' .

The eccentric toothed wheels 36 and 37 are joined together by an inner race member 60 of a bearing race having an annular array of rollers 62 and an outer race member 64. The bearing race is centred on the actual centres of the eccentric toothed wheels 36 and 37, i.e. actual centres 42 and 44 in Figure 4.

Similarly, eccentric wheels 38 and 39 are joined together by the inner race member 66 of a similar bearing race, the outer member of which is denoted by reference numeral 68, and having an annular array of rollers denoted by reference numeral 67.

Around the outer race members of both bearings is a connecting member 70 which secures the two pairs of toothed wheels 36, 37 and 38, 39 in perfect meshing alignment. Since the centres of the bearing assemblies 64 and 68 are concentric to the actual centres of the circular toothed wheels 36, 37 (and 38, 39) respec ively, the net effect is that the actual centres of the toothed wheels 36, 37 and 38, 39 remain at a fixed distance apart whilst the wheels rotate eccentrically around their centres of rotation 34, 40 etc. respectively.

Since the bearing rings 60 and 66 and 60' and 66' are hollow, the shafts 34, 34', 40 and 40' can pass inside the perimeter of said bearing rings.

For clarity, the connecting member 70 and 70' is in part cut away so as to reveal in crosssection the detail of the bearing shells such as 64 and 60 and the roller bearings or pins 62, and the connecting member 70 therearound.

It will be appreciated that just as the toothed wheels 36, 37 and 38, 39 are circular members rotating about axis other than their own geometric centres, the circular bearing rings 60, 66 also rotate about axes other than their geometric centres. The bearing rings 60, 66 interact, via the rollers 62, 67 and the outer races 64, 68, with the connecting member 70, to keep the toothed wheels 36, 37 and 38, 39 in meshing contact. Similar, but

concentrically mounted, bearings interact with connecting members 61, 62 to keep the gears 56, 57 and 58, 59 in meshing contact. Thus as the drive mechanism rotates, meshing contact is maintained but the eccentric mounting of toothed wheels 36, 37, 38, 39 causes the drive shaft 40 to move. Since the gears 56, 57, 58, 59 are concentrically mounted, the movement of the drive shaft 40 is contained to an arc centred on the axis 52 of the rotary output shaft 54.

The connecting members 61, -70 are shown in dotted outline in both Figures 3 and 4 and are labelled as such.

In the right hand half of Figure 5, the corresponding structures function in a corresponding manner.

«

Since the connecting members 61, 63, 70 determine the path of movement of the shaft 40, the slide bearings are not necessary in the embodiment of Figure 5. However, for convenience they are shown in Figure 5, to clarify the details of their construction.

The left hand end slide bearing made up of upper and lower bearing members 48 and 46 is shown in Figure 5 sectioned through the centre of the shaft 40, thus revealing the interior of the slideway 50 and the lower bearing member 46. The slideway is shown as comprising an aperture in member 72.

A similar slideway is, of course, provided at the opposite end of the shaft 40 in a member 74 and this slideway is denoted by reference to numeral 51. Here it is only shown in dotted outline.

Similar members 72' and 74' provide slideways 50' and 51 for the shaft 40 ' .

Aligned with but below the shaft 40 can be seen the out shaft 54. However, in the view of Figure 5, the toothe wheels 58, 59, 58' and 59' are immediately below the toothed wheels 56, 57, 56' and 57' and consequently ho of the former can be seen in Figure 5, since they are completely hidden by the toothed wheels 56, 56', 57 and 57' which are keyed to and rotate with the shafts 40 an 40" respectively.

In the same way as toothed wheel pairs 36, 38 (37, 39) held together so that the actual centres of the meshing gears 36, 38 etc are .kept at a constant distance apart, the meshing gears 56, 58; 57 59; 56 ^« , 58' and 57', 59* also kept similarly at a constant distance apart whilst constant meshing engagement, by shells and bearing assemblies generally designated 61, 63, 61' and 63' whi units are similar to the units 70, 70' associated with gears 36, 38 etc.

Balancing weights are shown at 76 and 78 and at 80 and

Similar balancing weights may be required on the shaft and in association with the toothed wheel 56 and the toothed wheel 56'.

As will be seen from Figure 5, the two trains of toothe wheels driven by the shafts 40 and 40' respectively ar identical but are arranged 180° of out phase so that wh the throw of the eccentric gears 38 and 39 is to one ha the throw due to the eccentricity of the other eccentri gears 38' and 39' , is to the opposite hand. Figure 5 shows the position of the eccentrically mounted pairs o

gears in which the gears 38 and 39 in the left hand train have been displaced to the maximum amount away from the driving shafts 26 and 18, whilst the other pair 38' and 39' , are displaced to their maximum amount in the opposite direction towards the shafts 26 and 18.

The operation of the transmission unit is to link ( in manner known per se ) the pairs of rotary pistons in the cylinder via the output shaft 54, and because of the eccentrically mounted gear wheels, any speeding up of the shaft 26 is matched by a corresponding deceleration of the shaft 18 to points at which maximum and minimum speeds of rotation are achieved by shafts 26 and 18 respectively and thereafter deceleration of the shaft 26 is accompanied by a corresponding acceleration of the shaft 18 until the roles are reversed, whereupon the ^' process continues.

Figure 6 illustrates an alternative arrangement for the eccentrically mounted gears 36 and 38. In Figure 6 only a single pair of gears is employed (i.e. gears 37 and 39 are not required) and in order to obtain balance, two bearing assemblies are provided, one on each side of the toothed wheels .

For convenience the two toothed wheels are given different reference numerals from those in Figure 5 and are denoted by reference numerals 84 and 86 respectively. Toothed wheel 84 corresponds to wheel 36 and 86 to wheel 38. On each side of each toothed wheel extends a ring member, 88 and 90 in the case of toothed wheel 84 (and 92 and 94 in the case of toothed wheel 86). Each of the rings 88, 90, etc. comprises the inner race of a bearing assembly which includes needle rollers or the like (such as 96) and an outer race member 98 in the case of inner race member 88,

100 in the case of inner race 90 and 102 and 104 in the case of inner .race members 92 and 94.

Whereas 88, 90, 92 and 94 are firmly attached to the faces of their respective toothed wheels 84 and 86 respectively, the outer race members 98, 100, 102 and 104 are not secured to the toothed wheels 84 and 86 except insofar as they are connected through the bearings, thus the outer races 98, etc. can rotate freely relative to the inner races 88, etc. and two connecting members 106 and 108, which correspond in function to the single member 70 of Figure 5, can be made to connect outer race members 98 and 102 on the one hand and 100 and 104 on the other hand to ensure a continual meshing of the two gears 84 and 86.

The connecting member 106 is shown cut away and sectioned on the left hand side of Figure 6 to illustrate the detail of the inner and outer bearing races.

Although not shown, the rotary pistons of Figure 1 and the drive shafts and transmission system of Figure 5 modified in accordance with Figure 6 or otherwise, may be mounted within a single unitary housing comprising an internal combustion engine having a single output shaft 54.

It is possible that the concentric gear attached to the laterally displaceable shaft is better positioned centrally between the two eccentrically mounted toothed wheels for the purpose to achieve a more even load distribution on the slide bearings associated with the laterally displaceable shaft.

By using driven pinions such as 32, 32' having one half the number of teeth of driving pinions 30, 30', so four

strokes are obtained per revolution per rotary piston power unit.

If the driven pinions such as 32, 32' have only one quarter of the number of teeth of 30, 30' then one could obtain eight strokes per revolution per rotary piston power unit etc; in which case a different design of cylinder would be called for involving the use of a corresponding multiple of spark plugs and inlet and exhaust ports.

In Figure 7 an engine is shown in which a cylinder 110 containing the pistons is mounted between two housings 112 and 114 and the shaft 14 extends into housing 112 whilst shaft 18 extends into housing 114.

The toothed wheels 30, 30' are mounted thereon as previously described with reference to Figure 5 and the gear trains following on therefrom are the same as in Figure.5. except that only one pair of meshing gears 56, 58 is shown connecting to the output shaft 54 from the assembly in housing 112, and only one pair of meshing gears 56' , 58' is shown connecting to the output shaft 54 from the other assembly, in housing 114. Likewise, connecting members such as 61, 61' .are not shown.

The housing section 110 provides support at 116 for the bearings 118, 120 but terminates to permit the output shaft 52 to extend completely across from one bearing section 112 to the other 114.

Figure 8 shows schematically in plan a further transmission unit. The two sets of vanes drive respective toothed wheels 151, 153 through a shaft 155 and a sleeve

- 23 - shaft 157 respectively. The vane-bearing members and the cylinder are located to the left of the arrangement shown in Figure 8.

As in the previous arrangements, the toothed wheels 151, 153 drive respective pinions 159, 161 through the meshing of their teeth, and the pinions 159, 161 have half the number of teeth of the driving wheels 151, 153 so that four strokes are obtained per revolution in each chamber defined between any pair of vanes.

Each driven pinion 159, 161 is mounted on a shaft 160, 162 which also bears a respective first rotary unit 163, 165. Therefore, the toothed wheels 151, 153 will follow the first rotary members in their rotation, and accelerate and decelerate 180° out of phase during each, revolution of the engine.

Each first rotary unit 163, 165 is mounted in face-to-face opposed relationship with a respective second rotary unit 167, 169. Rotary movement is transmitted between each first rotary unit and its respective second rotary unit by respective transmission members (pins) 171, 173. The transmission pins 171, 173 are mounted in fixed 180° out of phase positions on the respective second rotary units 167, 169 and engage in slots 175, 177 in the respective first rotary units.

The two second rotary units 167, 169 are mounted on a common shaft, on which there is also mounted a gear 179. The gear 179 meshes with an output gear 181 mounted on a common output shaft 183.

In use, the output shaft 183 rotates at a substantially

constant velocity at any given moment (in that it is not subject to cyclical accelerations and decelerations), although it may accelerate or decelerate over any given long period of time as the operation of the engine is speeded up and slowed down. In the following discussion, rotation of the output shaft 183 will be taken as the constant reference velocity of rotation against which the rotation of other members is compared.

The two second rotary units 167, 169 are constrained to rotate at a constant velocity through the action of the gears 179, 181. The first and second rotary units of each pair are mounted face-to-face but their respective axes of rotation are 164, 166, 180 laterally displaced (offset). Accordingly, with the direction of offset shown in Figure 8, the respective transmission pins 171, 173 will slide in the respective slots 175, 177 of the first rotary units 163, 165 so as to be at the maximum distance from the axis of rotation of the respective first rotary units 163, 165 when the pins 171, 173 are at the bottom in Figure 8, and so as to be at the minimum distance from the axis of rotation of the respective first rotary units 163, 165 when the pins 171, 173 are at the top in Figure 8. The constant rotational speed of the second rotary units 167, 169, to which pins 171, 173 are fixed, will cause relatively rapid rotation of the first rotary units 163, 165 when each respective pin is close to the axis of rotation of each respective first rotary unit, and will cause relatively slow rotation of the first rotary units when each respective pin is remote from the axis of rotation of each respective first rotary unit. Accordingly, for each 360° rotation of the first and second rotary units, each first rotary unit 163, 165 will undergo one period of acceleration followed by one period

of deceleration.

Figures 9 and 10 show side views of the two first rotary units, showing the locations of the slots 175, 177 and the different positions adopted by the transmission pins 171, 173 as the rotary units rotate. The two second rotary units are shown in Figures 9 and 10 in broken outline, and it will be appreciated that the pins 171, 173 are constrained to follow the path of this outline so that their radial positions with respect to the first rotary units 163, 165 will vary as the assemblies rotate.

As can be seen in Figure 8, the two transmission pins 171, 173 fixed on the two second rotary units 167, 169 are arranged at 180° opposite to each other around ^" the second rotary units. Accordingly, the first rotary ^f unit 163 accelerates and the second rotary unit 165.decelerates at 180° out of phase with each other and vice versa. Through the actions of the pinions 159, 161 and toothed wheels 151, 153 the pattern of movement of the two vane-bearing members is constrained to follow the pattern of movement of the two first rotary units. Thus, the two vane-bearing members will each repeatedly accelerate and decelerate, so that the accelerations and decelerations are 180° out of phase between the two vane-bearing members.

The amount of the offset between the axes of rotation 164, 166 of the first rotary units 163, 165 and the axis of rotation 180 of the second rotary units 167, 169 determines the magnitude of the accelerations and decelerations of the vane-bearing members, and thus the compression ratio of the engine. Accordingly, the compression ratio of the -engine can be varied by varying the offset between the axes of rotation 164, 166, 180.

The subassembly comprising gear 179 rigidly connected to the two second rotary units 167, 169 on either side of it as seen in Figure 8, can be mounted so as to be able to slide within an arcuate channel (having the same centre as the output shaft 183). This will vary the lateral displacement between the axes of rotation of the first and second rotary units, thereby effectively altering the compression ratio of the engine, as explained above. This can be carried out as required and without interrupting the operation of the engine.

The transmission pins 171, 173 are mounted in sliders 200, which slide in the radial slots 175, 177. As shown in Figure 11, the slots 175, 177 have keyways 202 (omitted in Figures 9 and 10 for clarity) anQ the sliders 200 have complementary keyways. The keyways serve to restrain movement of the sliders axially of the first rotary units 163, 165, as such movement would result, if permitted, in the sliders 200 leaving the slots 175, 177.